U.S. patent number 5,520,008 [Application Number 08/296,572] was granted by the patent office on 1996-05-28 for centrifugal compressor and heat pump comprising.
This patent grant is currently assigned to I.D.E. Technologies Ltd.. Invention is credited to Arie Kanevski, Abraham Koren, David Olomutzki, Avraham Ophir.
United States Patent |
5,520,008 |
Ophir , et al. |
May 28, 1996 |
Centrifugal compressor and heat pump comprising
Abstract
A lightweight, heavy duty, large volume centrifugal compressor
for use in mechanical vapor compression systems, especially water
vapor compression systems in heat pump installations, said
compressor comprising a shaft driven propeller-like rotary member
consisting of a frusto-conical hub and a plurality of curved blades
made of a lightweight material, each being secured to said hub
along a longitudinal curved line and radially extending therefrom;
each pair of adjacent blades being interconnected by a bridging
membrane member of a lightweight material curvingly extending from
the roots of the leading edges of said adjacent blades to the tips
of the rear edges of the blades; said rotary member being
encompassed within a closely fitting shroud, so that curved vapor
flow channels are defined between each said pair of blades, their
associated membrane member, and the shroud. There is also provided
a mechanical water vapor compression heat pump system comprising a
pair of centrifugal compressors according to the invention
operating in series.
Inventors: |
Ophir; Avraham (Herzliya,
IL), Olomutzki; David (Kfar-Saba, IL),
Koren; Abraham (Holon, IL), Kanevski; Arie
(Kfar-Saba, IL) |
Assignee: |
I.D.E. Technologies Ltd.
(Ra'anana, IL)
|
Family
ID: |
11065256 |
Appl.
No.: |
08/296,572 |
Filed: |
August 26, 1994 |
Foreign Application Priority Data
Current U.S.
Class: |
62/268; 416/185;
416/188; 416/230; 62/324.6; 62/510 |
Current CPC
Class: |
F04D
29/023 (20130101); F04D 29/284 (20130101); F25B
1/053 (20130101); F25B 30/02 (20130101); F05D
2300/121 (20130101); F05D 2300/603 (20130101); F25B
43/043 (20130101) |
Current International
Class: |
F04D
29/00 (20060101); F04D 29/02 (20060101); F25B
1/04 (20060101); F25B 1/053 (20060101); F04D
29/28 (20060101); F25B 30/02 (20060101); F25B
30/00 (20060101); F04D 017/10 (); F25B
019/00 () |
Field of
Search: |
;416/185,188,230,186R
;62/268,100,118,467,306,510,324.6 |
References Cited
[Referenced By]
U.S. Patent Documents
Foreign Patent Documents
Primary Examiner: Bennett; Henry A.
Assistant Examiner: Doerrler; William C.
Attorney, Agent or Firm: Gerstenzang; William C. Sprung Horn
Kramer & Woods
Claims
We claim:
1. A lightweight, large volume centrifugal compressor for use in
mechanical vapor compression systems, especially water vapor
compression systems in heat pump installations, said compressor
being capable of handling a vapor flow rate of about 300-400
m.sup.3 /sec, providing a compression ratio of about 1:3 and
sustaining mechanical stresses such as occur at tip speeds of about
500 m/sec; said compressor comprising a propeller-like rotary
member consisting of a frusto-conical hub and a plurality of curved
blades made of a lightweight material, each being secured to said
hub along a longitudinal curved line and radially extending
therefrom; each pair of adjacent blades being interconnected by a
bridging membrane member of a lightweight material curvingly
extending from the roots of the leading edges (as defined herein)
of said adjacent blades to the tips of the rear edges of the blades
(as defined herein);
said rotary member being driven by a shaft passing through the
center of a stationary circular back plate bounding said rotary
member at the rear;
said rotary member being encompassed within a closely fitting
shroud, so that curved vapor flow channels are defined between each
said pair of blades, their associated membrane member, and the
shroud.
2. A compressor according to claim 1, wherein said hub is
manufactured of aluminum and said blades and said membrane members
are manufactured of a fiber-reinforced composite material.
3. A compressor according to claim 1, wherein said frusto-conical
hub is formed at its aft end with a co-axial frusto-conical recess
and is seated on a corresponding frusto-conical stationary support
cantilevered from said stationary back plate; said shaft driving
the frusto-conical hub passes through an axial bore in said
stationary support and rotates therein by the aid of a pair of
bearings located in said bore adjacent to its two ends; the center
of gravity of said rotary member being between said bearing
span.
4. A compressor according to claim 1, wherein each of said curved
blades is shaped so that the radius extending from the axis of the
hub to any point on the central line of the contour edge of the
blade is fully contained inside the blade.
5. A compressor according to claim 1, wherein there are provided
additional shorter blades (so-called "splitters") extending from
the aft end of said hub and terminating between each pair of
adjacent regular-length curved blades.
6. A mechanical water vapor compression heat pump system of the
type comprising an evaporator-freezer chamber, a compressor chamber
juxtaposed to said evaporator-freezer chamber and a condenser
chamber juxtaposed to said compressor chamber;
means for feeding water or an aqueous solution into said
evaporator-freezer chamber;
compressor means in said compressor chamber for reducing the
pressure in said evaporator-freezer chamber down to the water
triple point pressure to cause a portion of said water or aqueous
solution to vaporize and another portion to freeze;
said compressor means being further adapted to withdraw the vapor
produced within said evaporator-freezer chamber, transport it into
the compressor chamber, compressing it therein and transporting the
compressed vapor to said condenser chamber;
water spray means in said condenser chamber for cooling and
condensing said compressed vapor by direct heat exchange
therewith;
means to remove the condensate water together with the cooling
water from said condensor chamber;
vacuum pump means for evacuating non-condensible gases from said
condenser chamber and means for continuously removing ice-water
slurry from said evaporator-freezer chamber and circulating it
through heat exchanger means in a space to be cooled, located
outside said heat pump system; characterized in that:
said compressor means consist of a pair of centrifugal compressors
according to claim 1 operating in series and located at the
opposite ends of the compressor chamber which is designed as a
horizontal cylindrical vessel, each of said compressors being
designed as a complete sub-assembly with its adjacent end cover of
said compressor chamber; and
inter-cooling water spray means are provided in said compressor
vessel between said two compressors for cooling the vapor
compressed by the first stage compressor before it is further
compressed in the second stage compressor.
7. A heat pump system according to claim 6, wherein both the
evaporator-freezer chamber and the condenser chamber are juxtaposed
comparatively, closely to said compressor chamber and are connected
therewith by wide, comparatively short and curved vapor inlet and
outlet ducts, respectively, offering minimal resistance to the flow
of vapor from the freezer-evaporator to the compressor chamber and
of compressed vapor from the compressor chamber to the condenser
chamber.
8. A heat pump system according to claim 6, wherein said condenser
chamber is placed on top of said evaporator-freezer chamber both
forming together an integral unit, the bottom of the condenser
chamber serving as the top of the evaporator-freezer chamber.
Description
FIELD OF THE INVENTION
This invention is concerned with a large scale, high performance
heat pump installation operating on the principle of mechanical
water vapor compression. The invention also provides, for use in
the aforementioned heat pump installation, a novel, large volume
centrifugal compressor distinguished from and superior to
conventional compressors by virtue of its novel structural features
and its capacity to attain hitherto unachievable compression ratios
and vapor flow rates.
BACKGROUND OF THE INVENTION
Most conventional heat pumps, whether used for heating or cooling
purposes, utilize a refrigerant having suitable thermodynamic
properties such as ammonia or certain organic fluids, mainly
freons. Basically such heat pumps consist of a closed system
comprising an evaporator, a compressor, a condensor, if necessary
an expansion valve, and various controls. The working fluid
(refrigerant) evaporates in the evaporator at a low temperature and
pressure, extracting from the surroundings a quantity of heat equal
to its heat of vaporization. The refrigerant vapors are compressed
by the compressor to a pressure and temperature sufficiently high
to enable the refrigerant to condense in the condenser by giving up
its heat or vaporization to a stream of cooling water or to the
atmosphere.
Heat pumps using water as the refrigerant have also been proposed
(see for example U.S. Pat. No. 4,003,213 and Israel Patent 64871)
and such systems include ejectors, absorption systems and
mechanical vapor compression (MVC) systems. The use of water as a
refrigerant is thermodynamically desirable owing to its good
thermophysical properties and the advantages of employing direct
contact heat transfer, eliminating the need for costly and
thermodynamically inefficient heat exchangers. Furthermore, water
is the most "environmentally friendly" working fluid available, in
contrast with currently used organic working fluids (CFCs) which
are environmentally damaging and are likely to be restricted or
banned altogether in the coming decade.
Known heat pumps employing water as a working fluid of the ejector
and absorption system types are characterized by low efficiency,
whereas MVC systems have a much higher efficiency, typically about
2 to 3 times greater. However, a major difficulty involved in the
use of water as a refrigerant in MVC systems is the very high
specific volume of water vapor which requires the use of a very
large compressor. Thus, in a large size refrigeration heat pump
having a cooling capacity of about 3 to 10 MW, the required flow
rate of water vapor would be about 300-400 m.sup.3 /sec which is
considered a relatively high volumetric flow rate. In addition, for
a 20.degree.-30.degree. C. temperature difference (between the
space to be cooled and the ambient temperature of the air or
cooling water) a compression ratio (CR) of the order of 1:9 would
be required.
For this range of flowrates and compression ratios, two basic
compressor types are suitable, namely axial and centrifugal. The
axial type, as used mainly on aircraft engines is well developed
and has a high efficiency but is expensive to produce, therefore
the centrifugal type is the most promising for this application.
However, to date no centrifugal compressors have been developed
which come even close to fulfilling the target specification (i.e.
300-400 m.sup.3 /sec., 1:9 CR) mentioned above. Compression ratio
is a function of tip speed. Typical tip speeds found on small
aluminum compressors are of the order of 500 m/sec which gives a CR
of approximately 1:3.
Conventional large diameter compressors, which in general do not go
beyond 1.6-1.7 m impeller diameter are mainly made of fabricated
steel construction (Aluminum alloy casting or machining from solid
as used on smaller machines, is not practical in the larger sizes
due to difficulties in cooling massive metal sections). Fabricated
steel construction generally involves welding individual cast steel
blades onto a solid cone (e.g. see Allis Chalmers, Catalogue, 1980,
page 337). Such designs are not capable of sustaining the
mechanical loads found at say 500 m/sec tip speed due to stress
limitations on welded sections. Hence the tip speeds attained by
such designs are generally quite low, resulting in compression
ratios not more than approximately 1:1.6. This severely restricts
the range of process applications. In addition, the sheer weight of
the rotor resulting from such construction methods entails a
complicated and expensive rotor support system.
Thus, to summarize, conventional large centrifugal compressors
exhibit limited CR and volumetric capacity and are costly to
manufacture.
OBJECT OF THE INVENTION
It is the object of the present invention to provide an
economically feasible large scale heat pump installation operating
on the principle of mechanical water vapor compression.
It is a further object of the invention to provide a large volume
centrifugal compressor for use in a water vapor compression heat
pump installation, having high compression ratios of about 1:3 at
tip speeds of about 500 m/sec.
SUMMARY OF THE INVENTION
In accordance with one aspect of the invention there is provided a
mechanical water vapor compression heat pump system of the type
comprising an evaporator-freezer chamber, a compressor chamber
juxtaposed to said evaporator-freezer chamber and a condenser
chamber juxtaposed to said compressor chamber;
means for feeding water or an aqueous solution into said
evaporator-freezer chamber;
compressor means in said compressor chamber for reducing the
pressure in said evaporator-freezer chamber down to the water
triple point pressure to cause a portion of said water or aqueous
solution to vaporize and another portion to freeze;
said compressor means being further adapted to withdraw the vapor
produced within said evaporator-freezer chamber, transport it into
the compressor chamber, compressing it therein and transporting the
compressed vapor to said condenser chamber;
water spray means in said condenser chamber for cooling and
condensing said compressed vapor by direct heat exchange
therewith;
means to remove the condensate water together with the cooling
water from said condenser chamber;
vacuum pump means for evacuating non-condensible gases from said
condenser chamber and means for continuously removing ice-water
slurry from said evaporator-freezer chamber and circulating it
through heat exchanger means in a space to be cooled, located
outside said heat pump system; characterized in that:
said compressor means consist of a pair of centrifugal compressors
according to the invention (as defined hereinafter) operating in
series and located at the opposite ends of the compressor chamber
which is designed as a horizontal cylindrical vessel, each of said
compressors being designed as a complete sub-assembly with its
adjacent end cover of said compressor chamber; and
inter-cooling water spray means are provided in said compressor
vessel between said two compressors for cooling the vapor
compressed by the first stage compressor before it is further
compressed in the second stage compressor.
In accordance with another preferred embodiment of the invention,
both the evaporator-freezer chamber and the condensor chamber are
juxtaposed comparatively closely to said compressor chamber and are
connected therewith by wide, comparatively short and curved vapor
inlet and outlet ducts, respectively, offering minimal resistance
to the flow of vapor from the freezer-evaporator to the compressor
chamber and of compressed vapor from the compressor chamber to the
condensor chamber. This eliminates the use of complicated ducting
and transfer passages thus giving rise to savings in frictional
losses and, more importantly, helping to preserve uniform velocity
profiles at the compressor inlet sections.
In accordance with yet another, most preferred embodiment of the
invention said condensing chamber is placed on top of said
evaporator-freezer chamber both forming together an integral unit,
the bottom of the condensing chamber serving as the top of the
evaporator-freezer chamber and being subjected to only very low
pressure differences between both its sides.
In accordance with another aspect of the invention, there is
provided a lightweight, large volume centrifugal compressor for use
in mechanical vapor compression systems, especially water vapor
compression systems in a heat pump installations, said compressor
being capable of handling a vapor flow rate of about 300-400
m.sup.3 /sec, providing a compression ratio of about 1:3 and
sustaining mechanical stresses such as occur at tip speeds of about
500 m/see; said compressor comprising a propeller-like rotary
member consisting of a frusto-conical hub and a plurality of curved
blades made of a lightweight material, each being secured to said
hub along a longitudinal curved line and radially extending
therefrom; each pair of adjacent blades being interconnected by a
bridging membrane member of a lightweight material curvingly
extending from the roots of the leading edges (as defined further
on) of said adjacent blades to the tips of the rear edges of the
blades (as defined hereinbelow);
said rotary member being driven by a shaft passing through the
center of a stationary circular back plate bounding said rotary
member at the rear;
said rotary member being encompassed within a closely fitting
shroud, so that curved vapor flow channels are defined between each
said pair of blades, their associated membrane member, and the
shroud.
It should be noted that dead spaces are defined in the compressor
between the back plate, the hub, the adjacent blades and the
membrane members, thus significantly reducing the weight of the
rotary member which results in reducing mechanical stresses in the
rotary member and enables to achieve a higher tip speed and
consequently higher compression ratios.
Said hub is preferably manufactured of aluminum and said blades and
said membrane members are preferably manufactured of a composite
material thus significantly reducing the weight of the rotary
member, which also results in reducing mechanical stresses in the
rotary member and enables to achieve higher tip speed and,
consequently, higher compression ratio.
Several terms have to be defined at this stage in order to simplify
the further description of the rotary member .sub.--. The smaller
end of the frusto-conical hub .sub.-- will be referred to as
"forward end"; and the larger end of the hub .sub.-- as its "aft
end". The edges of the blades .sub.-- are termed (see FIG. 4) as
follows: A--the blade root; B--the leading edge; C--the contour
edge; D--the trailing edge; and E--the rear edge.
In a preferred embodiment of the invention said frusto-conical hub
is formed at its aft end with a co-axial frusto-conical recess and
is seated on a corresponding frusto-conical stationary support
cantilevered from said stationary back plate; said shaft driving
the frusto-conical hub passes through an axial bore in said
stationary support and rotates therein by the aid of a pair of
bearings located in said bore adjacent to its two ends; the center
of gravity of said rotary member being between said bearing
span.
This embodiment (a) allows further reduction of the weight of the
rotary member owing to the recess; (b) shortens the bending arm
(moment) on the shaft, thus allowing a reduction in shaft diameter,
due to locating both the stationary support and the pair of
bearings inside the rotary member.
In a preferred embodiment of the invention each curved blade is
shaped so that the radius extending from the axis of the hub to any
point on the central line of the contour edge of the blade is
full), contained inside the blade. Such a construction practically
eliminates bending forces on the blades, allowing the centrifugal
forces to pull the blades only in the radial direction. This
permits the structural fiber reinforced (composite) material to
operate under favorable mechanical conditions, i.e., direct
tension. This maximizes the permissible tip speed limit.
In one preferred embodiment of the invention there are provided
additional shorter blades (so-called "splitters") extending from
the aft end of said hub and terminating between each pair of
adjacent regular-length curved blades.
BRIEF DESCRIPTION OF THE DRAWINGS
The invention will now be further described in more detail with the
aid of the accompanying non-limiting drawings, in which:
FIG. 1 is a schematic perspective view of a typical heat pump
installation according to one embodiment of the invention;
FIG. 2a is a schematic cross-sectional view of the heat pump
installation of FIG. 1;
FIG. 2b is a schematic top view of the heat pump installation of
FIG. 1;
FIG. 3a is a schematic axial cross-sectional view of the compressor
vessel of the heat pump installation of FIG. 1 taken along line
A--A in FIG. 2b;
FIG. 3b is a schematic cross-sectional view of the
evaporator-freezer and the condenser units of the heat pump
installation according to FIG. 1 taken along the lines B--B in FIG.
2a;
FIG. 4 is an axial cross-section of a compressor according to the
invention;
FIG. 5 is a radial cross-section of the rotary member along lines
V--V in FIG. 4; and
FIG. 6 is a schematic axial view of the rotary member from the
forward end, showing only one pair of opposing blades.
DETAILED DESCRIPTION OF THE INVENTION
The heat pump installation
As shown schematically in FIGS. 1, 2a and 2b a mechanical water
vapor compression heat pump installation generally referenced 1,
according to one embodiment of the invention, comprises an
evaporator-freezer unit (or flash chamber) 2, connected by means of
a vapor inlet duct 3 to an adjacent cylindrical compressor vessel 4
which, in turn, is connected by means of a compressed vapor duct 5
to a condenser chamber 6 located above the evaporator-freezer 2 and
integral therewith.
The feed water enters the heat pump installation via the
evaporator-freezer 2 which is maintained at vacuum conditions by a
pair of compressors 7, 7' operating in series and located at
opposite ends of the cylindrical compressor vessel 4. The water in
the evaporator-freezer 2 is thereby cooled by evaporation to the
water triple point (about 0.degree. C. and 4.6 mm/Hg). The
evaporator-freezer 2 is provided with an agitator 8 with scoops,
driven by an external motor, designed to continuously agitate the
ice/water slurry in the evaporator-freezer 2, the surface layer of
which is thus constantly renewed, preventing the build-up of a
stagnant ice layer and maximizing the coefficient of heat transfer
(by direct evaporation). In addition, the scoops of the agitator 8
are designed to continuously wet the walls of the
evaporator-freezer chamber 2 in order to prevent the formation of
"chunk" ice and to promote the formation of discrete ice crystals.
This is important in order to avoid eventual blockage of the exit
to the evaporator-freezer 2 by ice formation. Alternatively, or in
addition, the formation of ice in small crystal form may also be
assisted by adding salt to the feed water.
The vapor produced in the evaporator-freezer 2 passes through the
vapor inlet duct 3 into the compressor chamber 4 at 0.degree. C.
and is compressed therein by the first stage compressor 7 at a
compression ratio of about 1:3. The compressed vapor is directed by
the aerodynamic flow channels formed by the compressor shroud 9 (as
explained hereinbelow) backwards in the axial direction of the
compressor chamber 4 towards the second stage compressor 7' and its
associated shroud 9', wherein it is further compressed by the same
ratio of approximately 1:3, so that the total compression ratio of
the vapor is approximately 1:9. Between the first and second stage
compressors 7 and 7' there is interposed a direct water injection
de-superheater (or intercooler) 41 which brings the inlet
temperature of the vapor into the second stage compressor 7' down
to about 15.degree. C. Between the de-superheater 41 and the second
stage compressor 7' there is interposed a conventional droplet
separator 42. The vapor exiting the second stage compressor 7' has
a saturation temperature exceeding ambient temperature or that of
available cooling water, thus permitting heat rejection.
The compressed vapor is passed from the second stage compressor 7'
into the condenser unit 6 consisting of a packed bed provided with
cooling water spray means 61 at the top, fed by a water circulation
pump. The compressed water vapor rises in the condenser 6 through
the packed bed here it comes into direct counter-current contact
with the downward flowing cooling water. The vapor condenses and
the latent heat of condensation absorbed by the cooling water is
rejected to the atmosphere via the condensate and cooling water
which are removed together from the system. The condenser 6 is
continuously purged of non-condensible gases by a vacuum pump via
the duct 62 (FIG. 3b).
It should be noted that the circulating pump providing the cooling
water 2 to the condenser 6 need only supply enough head to overcome
frictional losses, since the major part of the head required to
lift the cooling water up to the top of the condenser 6 is supplied
by the vacuum in the system.
The water/ice slurry produced in the evaporator-freezer 2 can be
conveniently pumped out, concentrated if desired and delivered to
the end-user, i.e. the space which is to be cooled by the heat pump
installation.
It can be seen from the abovementioned figures that the total
layout of the installation is very compact, with the two
compressors 7, 7' facing each other at either end of the compressor
vessel 4. For flexibility of operation, each compressor 7 and 7' is
driven independently by an externally mounted, frequency converter
controlled electric motor 43, 43'. The diffusers are arranged to
turn axially, thus facilitating the flow of vapor from the exit of
the first stage via the de-superheater 41 and droplet separator 42
to the intake of the second stage. By placing both compressors
within the compressor vessel 4, considerable economies are achieved
in that the compressor shrouds 9, 9' can be constructed from very
light materials since they do not have to withstand the full force
of vacuum (approximately 700-750 mm/Hg) which force is taken up by
the pressure vessel walls. The shrouds 9, 9' thus only need to
withstand a pressure difference of at most 12 mm/Hg. On the other
hand, the compressor vessel 4 itself is designed in the shape of a
simple cylinder which is well capable of coping with the full force
of vacuum. Furthermore, the incorporation of both compressors 7, 7'
in the one compressor vessel 4 saves the cost of transfer piping
from the first stage compressor to the second stage compressor, as
in previously proposed installations.
The construction of the evaporator-freezer 2 and the condenser 6 is
an integral unit having a common partition which serves at the same
time as the bottom of the condenser 6 and the top of the evaporator
2, again saving some construction costs since the pressure
difference acting on this partition is only about 30-40 mm/Hg
instead of 750-755 mm/Hg which would result if the freezer top and
the condenser bottom were subjected to atmospheric pressure.
The compressor
FIG. 4 shows an axial cross-section of a compressor 10, which in
this particular embodiment comprises a rotary member 12, rotatable
around a frusto-conical stationary support 14. The compressor 10 is
surrounded by a curved annular shroud 16, and is bounded at the
rear by a stationary back plate 18, from which the stationary
support 14 is integrally cantilevered. The rotary member 12
consists of a frusto-conical hub 20 and a plurality of curved
blades 22, mounted on, and radially extending from the hub. The
design of the rotary member 12 is fundamentally lightweight, being
based on thin carbon fiber laminated shell type blades 22 connected
to a relatively small diameter hub 20 made of aluminum alloy.
In operation the vapor to be compressed enters the shroud 16
axially, passes through a plurality of aerodynamic channels, each
formed between the blades 22 and the shroud 16. The vapor is then
propelled away radially in a compressed condition from the annular
exit formed between the rear of the shroud 16 and the stationary
back plate 18.
The following novel elements of the compressor's construction were
developed by the applicant in order to minimize the weight of the
rotary member.
Each pair of the adjacent blades 22 are bridged by a monocoque
streamlined membrane member 32 (shown in axial cross-section of the
membrane 32 in FIG. 4; (a radial cross-section of the blades 22 and
the membrane members 32 is shown in FIG. 5). Each membrane 32
curvingly extends from the roots A of the leading edges B of
adjacent blades 22 to the tips of the rear edges E of these blades.
Due to this arrangement, vapor flow channels having a desired
aerodynamic shape are defined between each pair of adjacent blades
22, their associated membrane member 32, and the shroud 16. The
thin bridging membrane 32, forming the vapor channel floor, also
defines an empty space between it, the aluminum hub 14 and the back
plate 18. This entails considerable savings in weight, with
favorable implications on performance and cost. Conventional
compressors are designed with integral blades and hub, where the
aft diameter of the hub extends all the way to the trailing edges
of the blades. In this design, according to the invention the
maximum hub diameter (at its aft end) is considerably lower than
the maximum diameter of the blades which improves performance,
since the smaller the hub diameter, the lower the stresses produced
in it at a given speed.
The rotary member 12 is rotated by a shaft 24, one end of which is
splined to the hub 20, and its other end is coupled to a motor (not
shown). The combination of lightweight blades and membranes result
in lower stresses on the aluminum hub, which allows its center to
be hollowed out. As can be seen in FIG. 4, the aft end of the hub
20 is formed with a coaxial frusto-conical recess 25
correspondingly shaped so as to receive the stationary support 14
leaving a narrow gap between them. The stationary support 14, in
its turn, is provided with an axial bore 26, through which the
shaft 24 passes. The shaft 24 rotates on a pair of support bearings
28, positioned inside the stationary support 14 and located at both
ends of the bore 26. The hub 22 has at its forward end an
additional co-axial recess 30, wherein the end of shaft 24 is
accommodated. The recesses 26 and 30 further reduce the total
weight of the rotary member, which causes a further reduction in
mechanical stresses on the shaft and rotor support system. This
feature enables a relatively small diameter shaft and rotor support
to be used.
The rotary member 12 is designed and suspended by the bearings 28
in such a manner, that its center of gravity falls between the
bearings 28, rather than outside the bearings' span. Since this
results in a dynamically stiff system, a reduction in shaft
diameter is made possible.
As can be seen in FIG. 5, the blades are bonded and screwed to
metal brackets 36 which in turn are bolted to the aluminum hub 20.
The membrane 32, made of a carbon fiber laminate sheet which is
mechanically fastened to the sides of adjacent blades 22 defines
the flow channel "floor".
FIG. 6 is a schematic axial view of the rotary member 12 from the
forward end showing only a pair of opposing blades 22. It can be
seen, that the blades 22 are mounted onto the hub 20 along
longitudinal curved lines (see roots A of the blades 22 in FIG. 6).
One can also see, that the blades 22 extend radially from the hub
20, i.e., a radius R extending from the axis of the hub to any
point of the contour edge C (more exactly, to a point on its
central line) of the blade 22 will be fully contained inside the
blade. This construction leads to the following advantages:
The use of very thin lightweight flexible blades arranged in a
radial manner practically eliminates bending forces on the blades,
allowing the centrifugal forces to pull the blades only in the
radial direction, thus minimizing the total loads applied to the
rotary member i.e., this maximizes the permissible tip speed
limit.
* * * * *